Traction-drive transmission

ABSTRACT

The invention contemplates an improved mechanical traction drive providing a single control whereby a desired range of speed ratios between input and output shafts is selectively available, without requiring clutch connection to the power source, such as an internal-combustion engine. The transmission involves a combination of completely mechanical friction-roller and meshing-gear planetary systems which are connected at all times, whether the single control is operated to call for a forward drive, a stop or a reverse drive of the output shaft. For the particular disclosed forms, provision is made for automatic downshifting reaction to increasing load, and major axial-reaction forces are self-contained by rotating mechanism, thus (a) avoiding substantial axial force components on main-bearings of the mechanism, and (b) reducing to very small magnitudes the requisite force for selective control actuation.

This is a continuation-in-part of my copending application Ser. No.721,938, filed Sept. 10, 1976 now U.S. Pat. No. 4,098,146.

This invention relates to mechanical traction drives and in particularto those involving planetary systems of rolling and gear elements.

Structures of the character indicated which have thus far been proposedhave suffered from such mechanical complexity as to render themuncompetitive with more conventional (e.g., clutch-operated)transmissions. They suffer from excessive parts wear and heatdevelopment and are, in general, inadequate to the task of driving suchrelatively small vehicles as golf carts, garden tractors, snow plows andthe like. One improved mechanical traction drive of the characterindicated offering technical and economic advantages with respect to theconventional drives of today is disclosed in Dickinson application Ser.No. 614,606, filed Sept. 18, 1975 now abandoned.

It is a general object of the invention to provide further improvementin mechanical traction drives of the character indicated.

A specific object is to provide an improved smoothly variable mechanicaltraction drive, particularly for relatively small vehicles which aresubjected to frequent and massive changes in load, such as bull-dozinggarden tractors, snow plows, and the like, the drive being disengageableas for parking or towing.

Another specific object is to meet the above objects with a planetarysystem utilizing but a single control actuator to selectively determineany given speed within a finite range which includes forward drive, stopand reverse drive of the vehicle, without changing the running conditionof the drive motor or engine for the vehicle.

It is a further specific object to provide an improved transmissionutilizing a traction-roller planetary system wherein automatic overloadprotection is afforded by "downshift" of the drive ratio, all within apredetermined limiting condition, such as 90 percent for the coefficientof traction for the rolling elements, thereby avoiding "gross skid",assuring rolling-contact action at all times, and avoiding engine ormotor stall.

Still another specific object is to provide such a transmission whereinthe parts are balanced and spring loaded to minimize shift-controlforce, particularly for speeds near what I term a "hold" condition,meaning zero or substantially zero output speed, without disturbing thecontinuously connected and running condition of transmission parts;still further in such a transmission it is an object to selectivelyprovide total drive disconnection, using the same single shift-controlmeans to establish a true "neutral", as when parking with the enginerunning.

Another specific object is to provide a transmission meeting the aboveobjects and essentially confining the moving parts to a singlesubassembly, which is bodily removable from the transmission housing forready service and repair.

It is also a specific object to provide such a transmission whereinautomatic downshift is avoided throughout a range of zero tolight-torque loads for the forward-drive-selected condition of thetransmission, automatic downshift becoming operative only for greaterload-torque conditions; still further, in such a transmission, it is anobject to provide means whereby automatic downshift may be selectivelyavoided for the reverse-drive-selected condition of the transmission.

A general object is to meet the above objects with basically simplemechanism, of relatively great mechanical effectiveness, and of such lowcost as to be competitive with conventional transmissions of lessertechnical capability.

Other objects and various further features of the invention will bepointed out or will occur to those skilled in the art from a reading ofthe following specification in conjunction with the accompanyingdrawings. In said drawings, which show, for illustrative purposes only,a preferred form of the invention:

FIG. 1 is a partly broken-away longitudinal sectional view through atraction-drive transmission of the invention;

FIG. 1A is a simplified diagram to illustrate manually operatedcontrol-selection means for the transmission of FIG. 1;

FIG. 2 is a right-end view of the mechanism of FIG. 1 after itsright-end cover and output subassembly have been removed;

FIG. 3 is an exploded view in perspective, to show the planet-membercarrier of my transmission;

FIG. 4 is a left-end view of the mechanism of FIG. 1, with the housingpartly broken-away to reveal control parts;

FIG. 5 is a view in side elevation, to show control parts;

FIG. 6 is an exploded view in perspective of load-responsiveshift-control mechanism of FIG. 1, partly broken-away to reveal anoverall parts relationship;

FIGS. 7 and 8 are enlarged fragmentary views of shift mechanism of theinvention, taken respectively for the aspect of FIG. 4 and in plan view;

FIG. 9 is a graph to depict spring characteristics in use in themechanism of FIG. 1;

FIG. 10 is a graph to depict performance characteristics, namely outputtorque for various selected transmission ratios in the transmission ofFIG. 1;

FIG. 11 is a view similar to FIG. 8, to show the neutral-selectedcondition of modified shift mechanism of the invention; and

FIGS. 12 and 13 are simplified fragmentary views to respectively showparts relationships for forward-shifted and reverse-shifted departuresfrom the neutral condition of FIG. 11.

Referring initially to FIG. 1, the traction drive of the inventionaccepts continuous input rotation of a drive shaft 10 (as from aninternal-combustion engine, not shown) and converts the same to forwarddrive, "hold" (stop), or reverse drive of an output shaft 11, all inaccordance with the selective longitudinal positioning of a singlecontrol-rod element 12; the same control-rod element 12 is alsoselectively angularly shiftable to totally disengage the drive to a true"neutral", as for parking or towing conditions. The transmission will bedescribed in the context of illustrative use on a small vehicle such asa lawn or garden tractor equipped as for bulldozer, snow-plowing or thelike duty, for which the engine may be in the range of 10 to 20horsepower, but the principles of the invention will be understood to beof greater range of application. And remote actuation of control element12 will depend upon particular application convenience, with connectionto the outwardly exposed end of element 12; for example, in a manuallyoperated application, and externally projecting control arm 12' (FIG.1A) may be clamped to rod 12 beneath a suitably slotted cover plate 13',providing slightly offset "forward" and "reverse" longitudinal guideslots for arm 12' (offset δ₁), with a much greater offset (δ₂) betweenthem to permit approximately 30° rotational shift of rod 12, from "hold"to "neutral" selection, as will later become clear.

The transmission is contained in a relatively small cup-shaped housingbody 13, in the closed end of which the input shaft 10 is supported byan antifriction bearing 14 and is suitably sealed by means 15. Thecontrol element 12 is slidably supported through and sealed at 16 to anupper part of the closed end of body 13. The housing is closed by aremovable end bell 17 having a central hub in which the output shaft 11is shown supported by spaced needle and ball bearings 18-19 and issuitably sealed at 20. Shafts 10-11 include telescoping ends, withinterposed needle-bearing means 21; and a load is symbolized by anoutput bevel gear 22.

The Planetary Systems

Within the housing 13, and as part of a complete subassembly carried bythe input shaft 10, a single planet-element carrier 25 is rotatablymounted by bearing means 26 on shaft 10; the carrier 25 angularlypositions and carries plural planet rollers or wheels 27 and pluralplanet gears 28 in equal angularly spaced interlaced relation, therebyinterconnecting traction-roller and meshing-gear planetary systems, tobe described. Preferably, there are three planet rollers 27 and threeplanet gears 28. Each roller 27 has projecting rotary-support ends 29riding in needle bearings 30 in slide members 31, and members 31 are inturn guided by radial guide slots 32 in the carrier 25, to be more fullydescribed in connection with FIG. 3.

Each planet roller 27 is a single rigid element characterized by twolike rolling-contact surfaces 33-33' which are truncated-toroidal andconcave; surfaces 33-33' are sloped in generally axially-opposite andradially outward orientation, and the surfaces 33-33' may each be thesurface of revolution of a circular arc, about an axis outside thecircle from which the arc is taken.

The traction-roller planetary system comprises two like sun wheels35-35' mounted for independent and keyed axial sliding upon a drivesleeve 36, keyed at 37 to the input shaft 10; coupling means in the formof an axially flexible and torsionally stiff plate 34 is shown as themeans of establishing a keying connection from sleeve 36 to local keyrecesses 34' in the sun wheels 35-35'. The outer surfaces of wheels35-35' are convex and of opposed slope orientation, each beingpreferably the surface of revolution of a circular arc, of radius lessthan that of the circular arc defining the respective planet surfaces33-33'. Opposed Belleville washers or springs 38 are retained on sleeve36 by snap rings 39 to establish a predetermined axially squeezingpreload force of sun wheels 35-35' against the respective planetsurfaces 33-33', thus applying a radially outward force which tends tooutwardly displace the planet rollers 27. This displacement and forceare opposed by similar axially inward squeezing force applied to tworeaction rings 40-40', having antirotational support in housing 13. Suchsupport and the control and variation of squeezing-force action uponreaction rings 40-40' will be the subject of later and more detaileddiscussion, in connection with control by rod 12 and by load-responsivedownshift mechanism. It suffices here to note that the inwardly facingrolling-contact surfaces of reaction rings 40-40' may, like those of sunwheels 35-35', each be defined as a surface of revolution of a circulararc of radius less than that of the circular arc defining the pluralplanet surfaces 33-33'.

The meshing-gear planetary system comprises a drive or sun gear 41,keyed at 37 to shaft 10 and axially retained by and between snap rings42-43, along with sleeve 36, the inner rings of bearings 14-26, andaxial spacers 44-45 as appropriate. Gear 41 is in constant mesh with theplanet gears 28, and the latter are in constant mesh with the inwardlyfacing teeth of a ring gear 46 carried by a plate 46' that is keyed tothe output shaft; for total-drive disengagement, ring gear 46 is freelyrotatable on plate 46', and under drive conditions a clutch-dog rockerarm 23 pivotally mounted to plate 46' is spring-urged by means 24 toengage one of a plurality of dog slots in the adjacent end of ring gear46. Each planet gear 28 is seen in FIGS. 1, 2 and 3 to be needle-bearingmounted at 47 to a support pin 48 that is fixedly retained by means 49to part of the carrier structure 25.

The carrier 25 is seen in FIG. 3 to be essentially an assembly of aplanet-roller retainer casting 50 and a planet-gear cage subassembly 51,bolted together by means 52. Basically, the casting 50 is a continuousplate-like ring at its bearing-supported end 53, and formed withintegral arcuate angular segments 54 which extend axially and which areangularly spaced as needed for planet-roller clearance at the respectiveradially slotted guide locations 32. The planet-gear cage subassembly 51comprises annular plates 55-56, axially spaced and retained by spacingrivets 57. The plate 55 has three lobe-like projections to enablesecurely bolted fastening of three projections to the respective bodysegments 54; the plate 56 is circular and suitably bored at angularlyspaced locations for support of the three planet-gear pins 47. As bestseen in FIGS. 2 and 3, the aligned central openings 55'-56' of plates55-56 clear the teeth of gear 41.

Antirotational Support and Squeeze Control of Reaction Rings 49-41'

Primary reference is made to FIGS. 4, 5, and 6 in describing mechanismwhereby torsionally reactive antirotational support is provided for thereaction rings 40-40' and whereby a selectively applied controlpositioning of the shift rod 12 is subject to automatic load-responsivecorrection. Basically, the mechanism comprises (A) a torsionallyresilient suspension of an axially preloaded squeezing subassembly,shown in exploded array in the lower part of FIG. 6, and (B)cam-operated means referenced to the housing and associated with shiftrod 12, shown as a subassembly in exploded offset (on the alignment 60)from the squeezing subassembly. The shift axis of rod 12 is longitudinaland parallel to the central axis of the planetary systems (i.e., to theaxis of reaction rings 40-40'); and selected positioning of rod 12 istranslated, via cam means 61 and a cam-following arcuate yoke 62 havingpivoted reference at 63 to an axially fixed location in the housing,into corresponding changes in squeeze action. The yoke-pivot roll 63 isone of two, at substantially diametrically opposed locations on therespective arms of yoke 62, and a cam-follower roll 64 on a pedestal 64'at the midpoint of yoke 62 tracks shift-rod position via a guide or camslot 65 in means 61. The frame-reference for yoke pivots 63 will be seenin FIG. 2 to be provided by like diametrically opposed blocks 66,secured to housing 13, and each having an arcuate guide channel 67 inwhich the yoke pivot 63 is axially captive with a limited freedom forarcuate displacement.

The squeezing subassembly comprises spaced outer sleeve members 68-69 inaxial abutment with the respective reaction rings 40-40'; the radius atwhich sleeve 68 thus abuts reaction ring 40 is suggested by shadingbetween spaced arcuate phantom lines at 68' in FIG. 6, while the otherreacting ring 40' is seated within a locating skirt and against a bodyportion 69' of sleeve 69. Sleeve member 69 includes spaced brackets 70,recessed at 70', to engage and track the instantaneous axial position ofa first crank region 71 of each of the yoke arms, while a second crankregion 72 of each of the yoke arms is used for similar axial-positiontracking by the other sleeve member 69. Since regions 71-72 are onopposite sides of the axis of yoke-pivot means 63, the axialdisplacements of sleeves 68-69 are equal and opposite, in response toyoke actuation. Stiffly compliant Belleville-spring means 74 is reliedupon to apply a squeezing preload of sleeves 68-69 upon rings 40-49'; asshown, the outer radial limit of spring means 74 acts (to the right, inthe sense of FIGS. 1, 5 and 6) on sleeve 68, while a diametricallyextending beam 75 receives equal and opposite action from the innerlimit of spring means 74. Two tie rods 76 connect diametrically oppositeends of beam 75 to corresponding diametrically opposite locations on thesleeve body 69', so that spring action on beam 75 is directly translatedinto spring action (to the left, in the sense of FIGS. 1, 5 and 6) uponsleeve 69. Each connection of a tie rod 76 to an end of beam 75 is seenin FIG. 6 to involve a tie-rod guide member 77 which comprises alongitudinal channel to receive and locate the associated tie rod 76,the channel body being in turn located in an outwardly slotted end ofbeam 75; flanges or ears 78 on each member 77 bear against beam 75 atthe edges of each end slot thereof, and a washer 79 beneath the head ofeach tie rod 76 seats against the flanges 78 of the adjacent guidemember 77. Finally, a local recess 72' at a longitudinally centralregion of each guide member 77 coacts with yoke region 72 to respond toa shifted displacement of follower 64.

From the foregoing description of the squeezing subassembly, it will beunderstood that the instantaneous axial spacing of outer rings 40-40' isalways and solely a function of the instantaneous angular position ofyoke 62 about its pivot means 63. The force with which such spacing (ofrings 40-40') is held will be that which is needed to achieveequilibrium with the instantaneous radially outward displacement forceof planet rollers 27. The relatively great mechanical advantageattributable to the predominance of follower-crank radius R₁ overactuation-crank radii R₂ (see FIG. 5) means a correspondingly reducedreaction force as viewed along the displacement axis of control rod 12,but I prefer to select the force characteristic of spring means 74 so asto provide a "preload" force in such opposition to the radially outwarddisplacement force of planet rollers 27 that a nominal or "hold"position of yoke 62 is naturally retained. Thus, any adjusted shift ofrod 12 from its "hold" position will only involve differential actuationof the respective ends of the squeezing subassembly, so thatcontrol-force magnitudes can be kept at relatively very low levels,involving minimum reaction upon the housing or upon the controlmechanism.

A description of the squeezing subassembly is completed by noting thatboth tie rods 76 pass through aligned locating apertures in each of thereaction rings 40-40' and in the radial-plane wall of each of thebrackets 70, thus assuring angularly keyed integrity of all parts of thesubassembly. Additionally, the sleeve 69 is provided with diametricallyopposed pairs of angularly spaced arms 80; between each pair of arms 80,a compressionally preloaded spring 81 is seated on pads or washers 82. Asubstantial fraction of each pad 82 projects radially outside arms 80for torsionally resilient reacting engagement with adjacent side-wallregions of diametrically opposed local recesses 83 in housing member 17;these recesses may be seen in FIG. 4, but springs 81 have been omittedfrom FIG. 2 in order to permit viewing and identification of the guideblocks 66 for yoke-pivot action (already described).

The upper non-rotational control portion of FIG. 6 will be seen tocomprise a mounting plate 85 of sheet metal and establishing a guide forrod 12 with enlarged outer ends 86-86', to permit secure bolting, bymeans 90, to the interior of housing 13 (see also FIG. 2). Rod 12 has acircumferential groove for longitudinal-position tracking by the forkedend (92') of a pin 92 near the adjacent end member 61. A bracket arm 87is secured to and projects from the mounting-plate enlargement 86,providing extended frame reference for a slide block 88; block 88 has adovetail-guide (89) relation to arm 87, being adjustably positionablethrough selected setting of a lead screw as suggested at 91 in FIG. 2.The unpinned end 93 of control plate 61 is arcuately contoured forguidance between spaced shoulders 94 of a transverse groove in block 88,and capping plates 95 on the shouldered regions of block 88 sufficientlyoverlap the groove for captive retention of end 93 of the control member61. Thus, control member 61 and its cam 65 span a range of angularpositions of follower 64 about shafts 10-11, and throughout this range,cam-follower roll 64 is in constant tracking engagement with the cam 65.

It will be seen (FIG. 8) that an axial shifting of rod 12 will cause pin92 to pivotally displace cam plate 61 about an instantaneous center 93'of the rounded end 93 of plate 61, at a frame-reference longitudinallocation determined by the setting at 91, for block 88 in its guide 89.Such displacement of plate 61 will change the instantaneous location ofcam-follower (64) engagement along cam 65, thereby imparting arotational displacement of yoke 62 about its pivot means 63, and thusdirectly changing the axial spacing and, therefore the squeezing actionof reaction rings 40-40', as well as the preloaded condition of springmeans 74. As will later be more clear, the position of block 88 ispreferably set at 91 such that the instantaneous center 93' is in thesame radial plane (about shafts 10-11) as is the effective center offollower roll 64 when in "hold" position; and the effective longitudinallocation of shift arm 12' when in the wide slot between "hold" and"neutral" positions, is set such that the effective axis of cam 65 is inthis same radial plane (see phantom-line location 65' in FIG. 8). Thus,in "hold", no amount of load reaction will be effective to shift thefork 62 out of "hold" position.

The Belleville Springs And Their Loading

As already noted, each positional adjustment of the spacing of reactionwheels 40-40' is accompanied by a positional shift of the planet rollers27, in radial direction and extent, against the compressional preload ofthe sun wheels 35-35' (due to the combined effect of springs 38). Thespring means 74 merely relieves the net force encountered at control rod12; the characteristic and preload level of spring means 74 are selectedto substantially match or offset the instantaneous axial-force reactionfrom the preloading springs 38. In terms of control-rod (12)positioning, the net traction-drive ratio will always depend primarilyupon (a) the current positional setting of control rod 12 and (b) suchcorrective modification of the pivoted angle of yoke 62 as is achievedfor such setting by reason of load-reacting influence upon theantirotational springs 81 and the cam means 65-64. For the presentillustration in which forward, stop ("hold") and reverse drives areselectively available, such availability of "hold" (zero output speed)applies under load as well as under no-load conditions; the control-rodposition necessary to achieve "hold" will always be the same, but thecam follower 64 will assume various positions along the straight lengthof cam 65, depending upon the load condition. In any event, however, theabove-noted spring reaction, between inner-spring means 38 and thebalancing or offsetting effect of the outer-spring means 74, will alwaysbe operative upon the mechanism, and FIG. 9 is intended to assist in anappreciation of this point.

First, however, a preference is stated for the use of so-calledBelleville springs at 38-74 because they have the property of exhibitinga negative-rate coefficient for axial deflections beyond that deflectionat which their positive-rate coefficient ends. This positive-negativecharacter of the Belleville spring coefficient applies to such springsas are merely dished (frusto-conical) washers, as well as to suchsprings which are additionally characterized by radially slotted orother special features. Thus, the use of the expression "Belleville"herein is not to be understood as limiting the invention to plainfrusto-conical washers. And in the preferred employment of my invention,such springs 38 are under such preloaded condition as to assureoperation at all times in the negative-rate portion of their respectivecoefficients.

FIG. 9 is a simplified illustration of the use of the outer preloadedbalance spring means 74, in offsetting relation to the preloaded forcereaction from the sun-wheel spring means 38. The solid-line curve willbe seen to represent the characteristic of balance spring 74, with axialforce plotted in terms of increasing speed ratio, forward being taken aspositive. The selected usable range of spring means 38 is taken over thenegative-rate portion of the curve, between limits 100-101 of "Forward"and "Reverse" transmission, i.e., speed of output shaft 11 in relationto speed of input shaft 10; this speed ratio will be zero for the "Hold"condition 102. Parenthetic legends indicate that for greater speedratios in the forward direction, rings, 40-40' are squeezed moretogether (spring 74 being less compressed), and that for reduced speedratios including reverse, rings 40-40' are displaced more apart fromeach other (spring 74 being more compressed), all in accordance withshift yoke positions suggested by legend in FIG. 5. Thus, over the range101-100 depicted in FIG. 9, rings 40-40', most separated at 101, arebrought more together in the course of shifting through "Hold";throughout this direction of shifting through the range 101-100, spring74 is progressively expanded or decompressed, but in view of itsnegative-rate operation in this range, the preload force of spring 74increases. By the same token, shifting displacements which move rings40-40' more apart are accompanied by decreasing preload force of spring74.

For an assumed condition, in which "Hold" is the desired equilibrium(i.e., with shift rod 12 in its "Hold" position, and with moderate to noload-torque reaction at 22), the reaction characteristic from thepeloaded sun-wheel springs 38 (see dashed-line curve) is manifested as aspreading-displacement force between reaction rings 40-40', i.e.,counterclockwise reaction moment on yoke 62 in the sense of FIGS. 5 and6, and is seen to have been selected so as to cross or balance thesolid-line curve at the "Hold" point 102. With any shift-rod (12)displacement from "hold" and in the forward direction, a directiondifference force F₁ develops between the inner and outer spring systems(system 38 prevailing); the outer-ring squeezing which was accompaniedby development of this force F₁ institutes a radially inwarddisplacement of the planet rollers 27 and a spreading of sun wheels35-35' (with accompanying change in output speed ratio). And because ofthe indicated negative-rate nature of all spring systems, the requisiteplanet-wheel displacement response is quickly achieved with only mildlyincreased resistance. For shift-rod displacement from "Hold" and in the"Reverse" direction, a similar but oppositely-poled mild differenceforce F₂ develops (system 74 prevailing), to quickly accommodate theradially outwardly urged displacement of the planet rollers 27 (due toaccompanying axially approaching displacement of sun wheel 35-35').Thus, for any speed selection at 12, the maximum reaction forces F₁, F₂are of relatively small magnitude but are nevertheless of such polarityas to aid in returning the mechanism to the "Hold" position, subject ofcourse to the dragging resistance involved in restoring control rod toits "Hold" position, for a manually operated selection system, as viashift arm 12', I prefer that at least the forward guide slot shallinclude a rub strip or other means (schematically suggested by heavydashed line 58, in FIG. 1A) for enhancing this dragging resistance andthus retaining a given "Forward" selected position of arm 12'.

For many applications, the mechanism as described will be perfectlysatisfactory, but FIGS. 5 and 6 additionally illustrate a feature toprovide stronger resilient force to urge the shift yoke 62 and itsfollower 64 away from the extreme-forward displacement positions. Forsuch action, a mounting bracket 96 is fixed to sleeve 68 and retains theheaded end of an elongate square-section guide rod 97, keyed againstrotation by opposed walls of a radially slotted opening 98 in pedestal64'; and a compression spring 99 on rod 97 is preloaded against pedestal64', between bushings 99' and in accordance with preload adjustment ofthreaded means 97'. As yoke 62 is shifted to the right (forward speedselection), pedestal 64' is shifted to further compress spring 99, withthe result that spring 99 will tend to return yoke 62 to the neutral orto a less-than-full forward speed position.

Quite aside from the foregoing, it will be recalled that means externalto the traction drive may be provided to retain selection of the "Hold"drive-ratio position of shift rod 12, as suggested by the frame-basedslot system of FIG. 1A.

Operation of An Illustrative Transmission

For an illustrative embodiment wherein a 20 H.P. engine is to drive theshaft 10 for a range of FIG. 9 speed ratios and in which the limits 100and 101 are respectively in the order of +0.4:1 and -0.15:1, I havesuccessfully employed a 2:1 planetary gear ratio wherein sun and ringgears 41-46 are of 28 teeth and 56 teeth, respectively, and whereinthree 13-tooth planet gears 28 orbit on a 3.6-inch diameter circle. Atthe same time, the traction-roller planetary system has used planetaryrollers 27 in which the curvature radius of the concave rolling surfaceis 1.385 inches, a 40° arc of this radius being used to generate eachconcave surface of revolution, wherein the center for the curvatureradius is taken at 1.58-inch offset from the roller axis; to coact withthis planet-roll structure, each sun wheel and each reaction wheel has aconvex rolling-surface curvature of 1.066-inch radius, a 40' arc of thisradius being at 1-inch offset from the sun-wheel axis, and a 20° arc ofthis radius being at a 3.63-inch offset from the reaction-ring axis, togenerate the respective convex surfaces of revolution. The parts are runin a friction oil, an acceptable product being the Monsanto Co.synthetic hydrocarbon traction fluid commercially known as SANTOTRAC-50.

In the "Hold" setting of the shift yoke 62, and for the indicatedspecific embodiment, the effective radius of the orbiting circle ofplanet rollers 27 is such that the carrier 25 rotates at approximatelyone third of the speed of input shaft 10, and in the same direction asthe input shaft. For "Forward" output speeds (at shaft 11), carrier 25rotates at more than one third the speed of the input shaft, for exampleup to approximately two thirds of input shaft speed; for "reverse"output speeds, carrier 25 rotates at less than one third the speed ofthe input shaft, for example down to approximately one fourth of inputshaft speed. In every case, it is the effective instantaneous radius ofthe orbiting circle of the planet rolls which determines such carrierrotation, and it is the current pivoted orientation of yoke 62 whichdetermines the effective orbit-circle radius.

In "Hold", the cam plate 61 will be in the position shown at 61' in FIG.8, the shift rod 12 being pushed forward for forward-drive selection, asto the solid-line position of plate 61 in FIG. 8. Under a zero orsubstantially zero-load condition, the follower roll 64 is preferablypositioned, by torque-reaction springs 81, near that end of cam 65 whichis nearest pin 92 and, therefore, at substantial radial offset from thepivot center 93' of cam plate 61. Thus, for forward drive of a load,there will be a first direction of counter-rotational torque sustainedby housing 13 via springs 81 and the outer-ring squeezing assembly(lower part of FIG. 6), such counter-rotational torque involving angulardisplacement of said outer-ring squeezing assembly against springs 81;the maximum extent of this angular displacement is indicated α in FIG.7, is determined by the length of cam 65, and involves progressivelysmall pivot-angle offsets of the shift fork 62 (about its pivots 63)with respect to the "Hold" position thereof. For "Reverse"-selecteddrive of a load, there is a similar counter-rotational torque in theopposite direction, but cam follower 64 will be urged in the directionof the adjacent end of cam slot 65 and will therefore remain in its FIG.7 solid-line position throughout the lesser range of reverse-speed ratioselection; the lesser range β of reverse-drive selection, as comparedwith the greater range γ of forward-drive selection are shown by legendsin FIGS. 5 and 9.

Automatic downshifting and upshifting will be seen to be characteristicof the described mechanism, upon additional reference to FIG. 10. FIG.10 is illustrative of performance when the cam 65 of plate 61 slopesnegatively at an angle θ₁ of about 5 degrees (for the "Hold" condition)with respect to a radial plane of the central axis 10-11, said anglebeing identified in FIG. 8; also, for the FIG. 10 illustration, amaximum "forward" no-load ratio of 0.4 is attainable for a cam 65 slopeθ₂ of about 20 degrees, positive with respect to the same radial plane.

For any given "forward" selection, say the maximum ratio of 0.4:1, asbetween output speed (11) and input speed (10), increasing load ischaracterized by a "normal" speed reduction or droop due to internalslippage, until about 50 percent of maximum output torque is delivered,point A in FIG. 10, for an illustrative preload of springs 81. Point Ais on a first sloping alingment, to point A', which alignment determinesthe onset of load-torque reaction sufficient to commence displacement ofcam-follower 64 in its movement down cam 65. Such movement necessarilyentails a downshifting axial displacement of follower 64 and acorresponding rotation of its supporting yoke 62 about pivots 63. Theextent of downshift depends upon load-torque reaction, via springs 81,and accounts (along with normal slippage) for a more steep responsecharacteristic, from point A (about 500 in-lb. torque output) to point B(about 880 in-lb. torque output); with such downshifting, there isaccompanying speed reduction, from 0.4 to about 0.35 at point A, and toabout 0.1 at point B. Point B is on a second sloping alignment, to pointB', and is determined by the fact that cam-follower roll 64 has come tothe other end of cam 65 (near pivot center 93'). Further reaction toincreasing load torque can no longer be accommodated by cam 65 (for itsstated slope setting) and so performance will follow the alignment B-C,parallel to the initial slope from no-load to point A, until some partof the rolling system breaks traction.

Similar interpretations of FIG. 10 may be made for each of a pluralityof rod (12)-selected slope orientations of cam 65. Thus, for an initialno-load speed-ratio selection of 0.285:1, speed falls off at the normalslip rate until intersection at A" with the alignment A-A', whereuponload-torque reaction is operative to shift follower 64 in its path downcam 65. This continues until follower 64 abuts the low end of cam 65,denoted B", on the alignment B-B', at which point output torque is alittle more than 80 percent of maximum, and the transmission ratio hasbeen downshifted to approximately 0.08:1.

FIG. 10 is further of interest for its display of the straight-linecharacteristic which begins at a no-load transmission ratio of about0.09:1 and slopes at the normal speed-droop rate for its entire length,involving no change in direction at its intersections A"' with alignmentA-A" or B"' with alignment b-b". This characteristic will be understoodto apply for the phantom position 65' for cam 65 in FIG. 8, wherein cam65 is in a radial plane, i.e., perpendicular to the transmission axis.

The indicated course of speed reduction, for the 0.4:1 speed selectionand via points A and B (and for all lesser "Forward" selection ratios),will be seen to be within an outer envelope (solid-line curve) whichrepresents theoretical performance where 90 percent of the tractionlimit of rolling-contact parts is taken as the safety limit; in otherwords, rolling-contact traction may be achieved at higher loads on theseparts, but 90 percent provides an adequately safe margin upon which torely. The extension of solid-line curve A-B beyond point B would applyonly if the cam 65 were of greater length, it being recalled that theperformance curve proceeds in the direction C from point B becausefollower 64 intercepted the end of the particular cam 65 shown.

In similar fashion, FIG. 10 shows with another solid-line curve thefull-reverse selection available to the extent of a reverse-drive ratioof 0.146:1. From the no-load condition for this selection, the samenegative-slope "normal" speed droop is seen to occur for increasing loadtorques, to a point M (about 400 in-lbs.), beyond which the slope from Min the direction N completes a definition of the region within which theabove-mentioned 90 percent traction limit applies for the transmission;for example, M is the point at which this 90 percent limit is reachedfor maximum reverse-ratio selection, and M' is the corresponding limitachieved by selecting the lesser reverse ratio of 0.1:1.

"Hold" To "Neutral" Shift

Even though the transmission of my invention may be manually shifted to"Hold", i.e., to a condition wherein the transmission is running andfully connected to the output shaft 11, but wherein output shaftrotation is zero or substantially zero, further provision is made forattainment of a true "neutral", meaning total disconnection of thetransmission from the output shaft. In the form shown, this provision ismade at one or more dog-clutch arms 23, by clockwise rotation of thesame (in the sense of FIG. 1) to free arm 23 from dog engagement withring gear 46, thus severing the connection between ring gear 46 and itssupporting plate 46'. This disengagement is accomplished by theapproximately 30-degree rotation of rod 12 when selector arm 12' isangularly displaced across the span δ₂ of the "Hold"-"Neutral" slotregion in plate 13' (FIG. 1A), as will now be described.

At the rear end of the transmission, selector rod 12 is fitted with acrank arm 104 which is suitably formed with longitudinal and radialoffsets to clear ring gear 46 and the inner wall of end bell 17, for allpossible longitudinal and angular settings of rod 12. At its outer end,crank 104 has pin-and-slot engagement to a plate 105 in the form of ahorseshoe, spanning a diameter of a barrel-cam member 106, the arms ofthe horseshoe being connected to flexibly mounted lugs 107 of member 106at diametrically opposed locations. Cam member 106 is a collarlongitudinally guided on a cylindrical surface of a stem portion 108 ofend bell 17, and axially projecting cam-tooth formations 109 on member106 and the adjacent hub portion of end bell 17 are operative todetermine axial displacement of member 106, for angular shifts impartedby crank arm 104; in the course of such displacements away from theposition shown in FIG. 1, the horseshoe plate 105 flexes, and theadjacent end of the clutch rocker arm 23 is caused to displace in theclockwise sense, eventually disconnecting the engagement to ring gear 46when shift arm 12' is at or near the "Neutral" end of the displacementδ₂.

The disengaged condition will remain until selector arm 12' is returnedto its "Hold" position, allowing spring 24 to return the dog-clutchrocker arm 23 into engagement with the next available dog-slot formationin ring gear 46. In the course of driven rotation of gear 46 and itsplate 46', the dog-clutch rocker arm 23 is centrifugally urged to retainthe dogged engagement. In the event of two diametrically opposeddog-clutch devices 23 on plate 46', a balanced condition is inherent(with respect to the axis of shaft 11), but for the single clutchelement 23 shown, I have provided a suitable counterweight 23' fixed toplate 46'.

Modified Control Mechanism

FIGS. 11 to 13 illustrate a modified control mechanism wherein certainlimitations are inherently imposed upon the otherwise automaticallyoperative downshifting and upshifting functions described above, forcertain selected degrees of forward-drive and reverse-driverelationships. And the imposition of these limitations is shown for thecase of a control member or plate 161, fixedly pivoted at 120 to bracketmeans 121, secured as by mounting bolts 122 to part of the housingstructure of the transmission. Sector teeth 123 along an arc about pivot120 mesh with a rack 124 of revolution which forms part of the shift rod12 and which is the means of selecting desired angular orientation of acam slot 165 of control member 161. The cam-follower roll 64 is shown inslot 165 for the no-load limit of its travel about the drive axis of thetransmission.

The cam slot 165 is of dog-leg configuration, being characterized by afirst or downshift-control portion P and by a second ordownshift-relieving portion Q, at an angularly offset alignment withrespect to the alignment of the first portion P. For a purpose whichwill later become clear, the alignment of portion P is offset (to theextent D) from the axis of pivot 120; and for the "Neutral" conditionapplicable to the parts relation of FIG. 11, the cam portion P issubstantially in a single radial plane about the drive axis of thetransmission, so that whatever the load-torque and its tendency todisplace cam follower 64, there will be no cammed displacement byfollower 64, it being recalled that in FIG. 11, follower 64 is shown forthe no-load limit of its travel about the drive axis of thetransmission.

The angular offset of cam portion Q with respect to cam portion P isselected so that for the "full-high" or most-"Forward" orientation ofcontrol member 161 (as shown in solid lines in FIG. 12), the alignmentof cam portion Q is substantially in a single radial plane about thedrive axis of the transmission, it being noted that for this selecteddrive condition, the downshift-control portion P is at its greatestinclination with respect to any such radial plane. Also, the outer endof the downshift-relieving portion Q is selected to freely accommodatefollower 64 for the no-load limit of its travel about the drive axis,and for the most-"Forward" drive selection via rod 12, as shown in FIG.12. And the inner end of the downshift portion P is selected toaccommodate follower 64 for the full-load limit of its travel about thedrive axis, this latter limit being specifically identified at L_(f) inFIG. 12 as the critical ultimate point in space for the axis of followerroll 64 at the limit of its downshifting displacement. It will be notedthat the point L_(f) is at substantially the same angular orientation asthe location of pivot 120, both being viewed about the drive axis of thetransmission; and it will be further noted that by reason of the offsetD, the downshift-free zone Q becomes unavailable to follower 64 for"Neutral" and "Reverse"-selected orientations of control member 161.

For the described relationship of cam 165 and follower 64 and for themost-"Forward" selected orientation of control member 161 (FIG. 12),no-load or light-load conditions are accommodated without downshift fromthe "full-high" drive ratio, for a predetermined range of follower (64)displacement, as long as such displacement is within thedownshift-relieving or downshift-free zone represented by the effectivelength of cam portion Q. This fact enables smooth "high-gear" drive of avehicle, as along substantially level terrain. However, when load-torquereaction is sufficient to displace follower 64 beyond the indicatedrange, the cam portion P becomes operative to determine downshiftsettings as necessary to adapt to the instantaneous load torque.

For a "Reverse"-selected use of the embodiment of FIGS. 11 to 13, thereis no need or provision for a downshift-free zone of cam 165, and FIG.13 shows that for this selection follower 64 is at all times containedwithin the downshift zone P of cam 165. In fact, due to the above-notedoffset D of cam portion P from the pivot 120, a selected shift to the"Reverse" condition of FIG. 13 places the no-load limit of follower (64)travel at a location short of the juncture of zones P and Q; also, dueto this same offset D, the inner end of cam portion P determines adownshift-displacement limit L_(r) (for "Reverse"-drive selection) whichrepresents a more limited downshifting capability than does thedownshift-displacement limit L_(f) (for "Forward"-drive selection).

In certain tractor-drive employments of the invention, it is notnormally desirable that automatic downshift shall be applicable for a"Reverse"-drive selection. To a certain extent, the offset Daccommodates this objective by necessarily offering a reduced range ofdownshift displacement of follower 64, as compared to thedownshift-displacement range applicable for full "Forward" selection.However, in the embodiment of FIGS. 11 to 13, an even greater limitationis applicable to the range of available downshift displacement, in the"Reverse"-selected situation, by reason of a one-way engaging latch ordog element 125, pivotable about pin 120 and spring-urged by means 126to the position shown in all of FIGS. 11 to 13. In said position, thelower end of element 125 abuts the base of bracket 121, and the upperend is poised to prevent a downshifting displacement of follower 64 incam portion P, the latter function being available only for the"Reverse"-drive-selected orientation of control member 161 (FIG. 13).

If on the other hand, it should be desired to obtain a downshifted drivefor the "Reverse"-selected condition, a simple manual lever orpush-button operation will be understood to suffice for direct disablingactuation of element 125, as via a Bowden-wire cable 127, all asschematically suggested in FIG. 13 in application to the lower end ofelement 125. It will be further understood that, once downshifting hasoccurred in the "Reverse"-selected condition, the cable 127 may bereleased to permit spring 126 to urge element 125 (clockwise, in thesense of FIG. 13) against follower 64; thereafter, when load torquereduces sufficiently to enable follower 64 to assume its no-loadposition, element 125 will be automatically released by the excapingpassage of follower 64, allowing element 125 to return to its positionshown in FIG. 13.

While the invention has been described in detail for the forms shown, itwill be understood that modifications may be made without departure fromthe invention. For example, the cam 65 may be curved rather thanstraight, as may be appropriate for particular non-linear characterizingpurposes. Also, the means 91 will be understood to be suggestive ofvariation of the instantaneous pivot center of cam 61, whetherselectively fixed or continuously variable. And, for the offset-camembodiment of FIGS. 11 to 13, it will be understood that throughout thepredetermined low-load range of applicability of cam portion Q, there isnot only no automatic downshifting of the mechanism but there is also noreaction force to dislodge or disturb a "full-high" or most-"Forward"drive-selecting position of control rod 12, thus enabling more-free useof the operator's hands for steering and other purposes while in "fullhigh" and under light load torque.

What is claimed is:
 1. A transmission comprising a housing, a driveshaft and a driven shaft journaled for rotation in said housing on acommon axis, a variable-ratio coupling mechanism coupling said shafts,said mechanism including torque-responsive reaction means havingantirotational support about said axis in said housing, a control memberfor changing the coupling ratio between said input and output shafts,said torque-responsive reaction means including a cam pivotably carriedby said housing and a follower connected to said control member andhaving torque-responsive displaceability with respect to said cam formodifying the instantaneous control condition of said control member,and selectively operable means for varying the effective slopeorientation of said cam, whereby the rate at which the coupling ratio isvaried in response to load torque is a function of the instantaneousslope orientation of said cam.
 2. A transmission comprising a housing, adrive shaft and a driven shaft journaled for rotation in said housing ona common axis, a variable-ratio coupling mechanism coupling said shafts,said mechanism including reaction means having torsionally yieldableantirotational support about said axis in said housing, whereby inreaction to increasing torsional load said mechanism will be displacedin rotation about said axis and against increasing torsional resistanceby said yieldable means, said mechanism further including a controlmember for changing the coupling ratio between said input and outputshafts, cam and follower elements reacting to such displacement andincluding a cam carried by said housing and a follower connected to saidcontrol member for modifying the instantaneous control condition of saidcontrol member, and means for varying the effective slope orientation ofsaid cam, whereby the rate at which the coupling ratio is varied inresponse to load torque is a function of the instantaneous slopeorientation of said cam, said cam having at least one straight rampportion which is positioned in cooperative relation with said followerfor a full-high position of said control member, said one ramp portionhaving a directional axis which is in substantially a single radialplane and therefore normal to said common axis for said full-highposition of said control member, whereby a limited range of torsionalload may be developed for said full-high position without incurring achange in the coupling ratio between said input and output shafts. 3.The transmission of claim 2, in which said cam is pivotably mountedabout a pivot axis in a single radial plane, said cam having a secondstraight ramp portion inclined with respect to said first ramp portionand located between said first ramp portion and the pivot axis of saidcam, said control member having a neutral position wherein said secondramp portion is in substantially a radial plane and is therefore normalto said common axis.
 4. The transmission of claim 3, in which thealignment of said second ramp portion is so offset from alignment withthe pivot axis of said cam that for the neutral-selected position ofsaid control member, said first ramp portion is displaced out ofpossible follower engagement.
 5. The transmission of claim 3, in whichsaid cam is pivotably mounted about a pivot axis in a single radialplane, said cam having a second straight ramp portion inclined withrespect to said first ramp portion and located between said first rampportion and the pivot axis of said cam, said control member having areverse position wherein said second ramp portion is in substantially aradial plane and is therefore normal to said common axis.
 6. Thetransmission of claim 5, and one-way-engaging latch means in the path offollower movement for the reverse-selected position of said controlmember, said latch means being operative to retain said follower againstsubstantial movement along said second ram portion in the downshiftingdirection of reverse-selected operation.
 7. The transmission of claim 3,in which said control member is part of a control mechanism which alsoincludes a shift rod on an axis substantially parallel to said commonaxis, whereby for said limited range of torsional load there is nodevelopment of any force to change the selected position of said shiftrod.
 8. The transmission of claim 7, in which the pivot axis of said camis a fixed pivot axis and said cam includes sector-gear formations aboutsaid fixed pivot axis, said shift rod having rack formations engaged tosaid sector-gear formations.
 9. The transmission of claim 8, in whichsaid rack formations are teeth of revolution about the rod axis wherebysaid shift rod may be rotated for further control purposes withoutchange of a selected position of said control mechanism.
 10. Atransmission comprising a housing, a drive shaft and a driven shaftjournaled for rotation in said housing on a common axis, avariable-ratio coupling mechanism coupling said shafts, said mechanismincluding reaction means having torsionally yieldable antirotationalsupport about said axis in said housing, whereby in reaction toincreasing torsional load said mechanism will be displaced in rotationabout said axis and against increasing torsional resistance by saidyieldable means, said mechanism further including a control member forchanging the coupling ratio between said input and output shafts, camand follower elements reacting to such displacement and including a camcarried by said housing and a follower connected to said control memberfor modifying the instantaneous control condition of said controlmember, and means for varying the effective slope orientation of saidcam, whereby the rate at which the coupling ratio is varied in responseto load torque is a function of the instantaneous slope orientation ofsaid cam, said cam being pivotably mounted and having a ramp which isoffset from alignment with the pivotal axis of said cam, said ramphaving a finite follower-limiting end for torsionally loadeddownshifting displacement of said follower, and the direction of saidoffset being such that lower transmission-ratio accommodation of loadtorque is available for a forward-selected position of said controlmember than for a reverse-selected position thereof.
 11. Thetransmission of claim 1 or claim 10, in which said variable-ratiocoupling mechanism comprises an assembly including a carrier mounted onone of said shafts and constantly engaged rotatable elements on saidcarrier and on said shafts for variable-ratio coupling of said shafts,torsionally compliant means coactive between said housing and saidassembly for torsionally compliant angular displaceability of saidassembly as a function of load-torque reaction, axially displaceablereaction rings forming part of said assembly for varying theinput-output transmission ratio in accordance with the relative axialdisplacement of said rings, said follower being part of differentiallyoperative means for varying the axial separation of said rings and thusfor varying the input-output transmission ratio.
 12. The transmission ofclaim 6, in which said latch means includes selectively operable meansfor relieving the same from latching operability.